Compact indirect evaporative cooler

ABSTRACT

An indirect evaporative cooling system with a core greatly reduced in size compared to conventional evaporative cooling systems The system has a heat exchanger core having heat exchange plates defining a plurality of wet air flow passages and a plurality of dry air flow passages. At least one fan drives air through the passages. The dry air passages have a small height and a short length, configured so that a substantially laminar airflow having a raised shear rate arises in the dry air passages, and so that a back pressure across a length of the dry air passages remains low.

CROSS-REFERENCE TO RELATED APPLICATIONS

This application claims the benefit of U.S. Provisional PatentApplication No. 62/047,160 filed 8 Sep. 2014, which is incorporatedherein by reference.

TECHNICAL FIELD

The invention relates to an indirect evaporative cooler, and inparticular, an indirect evaporative cooler with a compact core. Thecompact core can also be used in a heat recovery heat exchanger.

BACKGROUND OF THE INVENTION

At its core, a modern indirect evaporative cooler typically comprises astack of thin parallel spaced-apart plates. The plates definealternating wet and dry air-flow passages. In the wet passages a“working” airstream passes over wetted surfaces, accepting and carryingaway sensible heat as well as latent heat of evaporation, leavingevaporatively cooled wet surfaces. In the dry passages an initially warmairstream is cooled as heat is transferred by convective transfer fromthe airstream to the cooled plate surfaces and by conductive heattransfer through the plates. The temperature gradient between theairstreams on either side of the thin plates drives the flow of heatfrom dry passage side to wet passage side.

A two-stage evaporative cooler improves over a single-stage cooler inthat an airstream pre-cooled in a first or “indirect” cooling stage isused as the working air stream in the wet passages in a second or“direct” stage. Alternatively, the cooled airstream is divided into twostreams, one of which is used to cool the living space, the other ofwhich is returned as “return air” to the wet passages. A pre-cooledworking air stream can carry away more heat from the wetted surfaces,further lowering the temperature of the evaporatively cooled wetsurfaces, thus lowering the temperature of the airstream beingconditioned in the dry passages.

In a heat-scavenging heat exchanger, on the other hand, thethermodynamics are worked in reverse. For example, when heating a livingspace in winter, cold outside air is drawn in through dry air channels.Warm humid air leaving the living space passes through wet channels incounter-current flow. As the outgoing warm humid air passes adjacent tothe cold incoming air, the water vapor in the outgoing air condenses,evolving heat of condensation (equal to the heat of vaporizationconsumed in cooling), and transferring sensible heat across the metalplates to warm the incoming air in the dry passages. As a result, a goodportion of the heat leaving the living space can be recaptured andreturned to the living space.

Indirect evaporative cooling systems offer many advantages overmechanically driven refrigerant-compression-type cooing systems, such aslower electrical consumption, higher reliability, and freedom fromenvironmentally detrimental refrigerants such as R-134.

However, although many improvements have been made over the years, amajor disadvantage in evaporative cooling systems is the considerablebulkiness of the heat exchanger, and particularly the core. The coolingcapacity of the evaporative cooler is generally considered to be afunction of the total evaporative surface housed in the system. Acommercial evaporative cooler can be several times the size of acomparable refrigerant-compression based cooling system.

Attempts have been made to improve the efficiency of heat exchangers. Asexplained in U.S. Pat. No. 5,718,848, the boundary layer of gas adjacentthe plates constitutes an obstruction to transfer of energy. Gases arenotorious insulators against conduction, and air is very resistant toheat transfer. According to U.S. Pat. No. 5,718,848, evaporation ofwater into the working air stream is improved by using multiple wickspositioned so that sufficient turbulence is developed to effect periodicrestart of the process of evaporation of moisture from the wicks.

A more recent proposal for improving heat exchanger efficiency isdescribed in U.S. Pat. No. 8,636,269. The evaporative heat exchanger isformed of corrugated sheets of material having one wettable surface andan opposed dry vapour-resistant surface. Sheets are stacked withwettable surfaces facing each other to form wet passages and drysurfaces facing each other to form dry passages between the sheets. Thegeneral direction of air flow is at an angle to the corrugations.Working air enters the labyrinthic matrix and encounters numerousintersections of the adjacent corrugated sheets. At each of theseintersections there is intense interaction between air and water,resulting in rapid evaporation of water from the wetted surfaces,thereby humidifying the working air and cooling the water on the wettedsurfaces. In the dry passages on the dry side of the sheets, hotincoming air to be conditioned loses thermal energy by interaction withthe evaporatively cooled sheet. According to U.S. Pat. No. 8,636,269,heat exchange between the wet and dry passages and evaporation withinthe wet passages can readily take place due to the intensity of mixingpromoted by the corrugated construction.

The conventional wisdom is thus to design the wet and dry air passagesfor increased turbulence in order to promote heat transfer andevaporative cooling. However, as turbulence increases, so doesresistance to flow, which can be measured as increased pressure dropover distance. This pressure drop must be overcome, for example, byprovision of additional or higher power air movers, e.g., fans, whichundesirably increases power consumption of the system. Adding fans alsoincreases construction and operating costs. It is known that pressuredrops can be mitigated by increasing the height of the airflow passages.However, considering that air is an insulator, as the height of airpassages is increased, the effectiveness of heat transfer is lowered.Further, an increase in plate spacing will increase the total height,and thus bulk, of the heat exchanger core.

It is disclosed in U.S. Pat. No. 8,636,269 that airflow at a shallowerangle to the direction of the corrugations results in smoother airflowand lower resistance, but at a penalty of reduced heat transferefficiency. The patent teaches that this loss of heat transferefficiency can be regained by extending the overall length of the core.Increasing passage length increases the bulk of the core. And sincepressure drop increases as the length of the air passages is increased,additional air movers will also be required. Thus, bulkiness appears tobe an inherent characteristic of heat exchangers.

US Patent Application No. 2004/0061245 (Maisotsenko et al) teaches anevaporative air cooling system with an improved heat exchange surfaceand an improved system of distributing evaporating fluids. The coolingsystem is comprised of plates having a water conducting layer on oneside and a low permeability layer on the other side. The low permeablelayer exhibits low conductivity to heat except when very thin and ispreferably plastic. Since heat transfer is good perpendicularly througha plastic layer but poor along the surface of a plastic layer, the“result of this differential heat transferability is that heat willtransfer from one side of the plate to the other along the interface ofthe plastic while at the same time heat will not readily transfer alongthe surface. The result is that discrete temperatures and a temperaturedifferential can occur at different points in the plate and it will notbe averaged due to the heat transfer by the plate.” (paras.[0026]-[0027]). Also mentioned is that when plate spacing is withincertain bracketed values (be 1.57 mm to 1.83 mm, 2.17 mm to 2.33 mm,2.16 to 2.87 mm, or 3.13 mm to 3.39 mm), the pressure drop across platesis reduced from 1% to 15%. Further, deposit build-up is reduced alongthe plate surfaces due to the transverse quarter wave, i.e., turbulence,increasing the dynamic energy of the flow in the direction of the flowat the boundary layer. Thus, Maisotsenko et at adhere to the acceptedwisdom that dynamic energy, i.e., turbulence, is needed in the air flowto break up the boundary layer for efficient heat transfer.

There remains a need for a more compact one or two stage evaporativecooler, as well as a compact heat exchanger for recapture of heat.

Any discussion of documents, acts, materials, devices, articles or thelike which has been included in the present specification is solely forthe purpose of providing a context for the present invention. It is notto be taken as an admission that any or all of these matters form partof the prior art base or were common general knowledge in the fieldrelevant to the present invention as it existed before the priority dateof each claim of this application.

Throughout this specification the word “comprise”, or variations such as“comprises” or “comprising”, will be understood to imply the inclusionof a stated element, integer or step, or group of elements, integers orsteps, but not the exclusion of any other element, integer or step, orgroup of elements, integers or steps.

In this specification, a statement that an element may be “at least oneof” a list of options is to be understood that the element may be anyone of the listed options, or may be any combination of two or more ofthe listed options.

SUMMARY OF THE INVENTION

According to a first aspect the present invention provides an indirectevaporative cooler comprising:

a heat exchanger core having heat exchange plates defining alternatingwet and dry air flow passages, wherein the relationship between thepassage height, measured as space between plates defining at least thedry air flow passages, and the length of the air flow passages fallswithin an area on a graph defined by the following points, whereinpassage length is plotted on one axis and passage height, measured asdistance between plates, is plotted on another axis:

Passage Passage Length (mm) Height (mm) 400 1.7-5.5 300 1.55-4.5  2001.3-3.5 100 0.9-2.7 75 0.8-2.3 50 0.7-1.7 25 0.6-1.1

According to a second aspect the present invention provides a method ofindirect evaporative cooling, the method comprising:

-   -   directing a laminar flow of air at a flow rate of 2.5-7.0 m/s        through at least dry air passages of a heat exchanger having        heat exchange plates defining alternating wet and dry air flow        passages, the plates having a separation of from 0.6 mm to 2.0        mm and defining air passages having a length of from 25 to 300        mm,    -   dividing the air after passing through the dry passages into        first and second air streams,    -   directing the first air stream into the wet air flow passages in        counter-current flow to the airflow in the dry passages, and    -   directing the second air stream to a space to be cooled.

According to a third aspect the present invention provides an indirectevaporative cooler comprising:

-   -   a heat exchanger core having heat exchange plates defining a        plurality of wet air flow passages and a plurality of dry air        flow passages; and    -   a fan configured to drive air through the dry air passages,    -   the dry air passages being configured so that a substantially        laminar airflow having a raised shear rate arises in the dry air        passages, and so that a back pressure across a length of the dry        air passages remains acceptably low.

In some embodiments of the invention, the relationship between thepassage height of at least the dry air flow passages and the length ofthe passages falls within an area on a graph defined by the followingpoints, wherein passage length is plotted on one axis and passage heightis plotted on another axis:

Passage Passage Length (mm) Height (mm) 500 2.8-4.5 400 2.5-3.7 3002.2-3.3 200 1.8-2.5 100 1.5-2.2 75 1.1-1.7 50 0.9-1.1 25 0.65-0.7 

In some embodiments of the invention the passage length is in the rangefrom 80 to 200 mm.

In some embodiments of the invention the space between plates definingdry passage height is from 0.6 mm to 2.5 mm, more preferably from 0.7 mmto 1.4 mm, and more preferably from 0.8 mm to 1.2 mm.

In some embodiments of the invention, the space between plates definingwet passage height is from 0.6 mm to 2.5 mm, more preferably from 0.7 mmto 1.4 mm, more preferably from 0.8 mm to 1.2 mm.

In some embodiments of the invention, air is flowed through the dry airflow passages at 3.0-7.0 m/s, more preferably at 2.5-4.0 m/s.

In some embodiments of the invention, in use the back pressure along thelength of the dry air passages is less than 100 pascal (Pa), preferablyless than 60 Pa, more preferably less than 50 Pa, most preferably lessthan 45 Pa.

Desiring a more compact heat exchanger core, and realizing that the“conventional wisdoms” of core design inevitably produce a large sizeindirect evaporative cooler core (i.e., using turbulence to break downthe insulating boundary layer and thus promote heat transfer; spacingplates far enough apart to prevent constriction of flow, etc.), thepresent inventors developed a design for a new type of core withsurprising compactness and efficiency.

The present invention recognises that when the spacing between theplates is greatly reduced and, at the same time, care is taken to ensurea substantially laminar flow rather than turbulent flow, goodevaporative cooling and good heat exchange occurs, allowing the lengthof the passages to be reduced. And, as the passage length is reduced,the pressure drop along this shorter passage length remains acceptable.

That is, by pushing air at high shear rates through passages havingnarrower spacing between plates—and shorter passages—the evaporativecooling system becomes not only compact in size, it also becomesefficient.

The advantages of the inventive design include:

-   -   good heat transfer since thinner boundary layers in the narrower        passages minimize the insulating effect of air,    -   more plate area for a given stack height since plate spacing is        reduced,    -   shorter passage lengths—as little as only 20% of the        conventional passage length—a major advantage in an installed        system, and    -   no flow impedance attributable to turbulence-induced flow        resistance, since flow is laminar.

The plates may be formed of any suitable material or materials,including metal and plastic, for example, PVC. The plates preferablyhave an absorbent flock on one side. The plates are preferably formedwith long arched sections between standoff spacers (see FIG. 6), with aradius of e.g. 10 mm, whereby static air in corners is minimized.

The invention has applications not only for indirect evaporativecoolers, but also to air-to-air heat exchangers (heat scavengers, heatrecovery systems) and/or water-to-air heat exchangers.

BRIEF DESCRIPTION OF THE DRAWINGS

An example of the invention will now be described with reference to theaccompanying drawings, in which like reference numbers indicate similarparts, and in which:

FIG. 1 shows a simplified core comprised of a stack of plates withalternating wet and dry passages;

FIG. 2 shows the CFD model, with variables being plate spacing or airpassage height (H) and passage length (L);

FIG. 3 shows the results of CFD in the range 0.6 to 1.5 mm passageheight and 25 mm to 100 mm passage length;

FIG. 4 shows the results of CFD in the range 0.6 to 5 mm passage heightand 25 to 500 mm passage length;

FIG. 5 is a graph showing inventive and preferred passage heights andlengths,

FIG. 6 is a close-up of a practical embodiment of a core structureshowing long arched sections between standoff spacers;

FIG. 7 schematically represents an evaporative cooler; and

FIG. 8 shows the design of an evaporative cooler according to oneembodiment of the invention.

DESCRIPTION OF THE PREFERRED EMBODIMENTS

FIG. 1 shows a known airflow configuration for a single core evaporativecooler system comprising parallel heat exchange plates 1. Each plate hasa wettable surface on one side and a dry vapour-resistant surface on theopposite side. Plates are stacked with wettable surfaces facing eachother to form wet air passages 2 and dry surfaces facing each other toform dry air passages 3 between the plates. Incoming source air 4 to becooled is directed through the dry passages 3 of heat exchanger 10. Uponexiting the dry passages, the dry cooled air stream is dividedapproximately evenly into a “supply air” stream 5 for cooling a livingspace and “return” or “working” air stream 6 which is directed into thewet passages 2. The wet passages 2 have a hydrophilic surface providedwith a wicking material 7 which is capable of being kept continuouslywet by being intermittently or continuously charged with water. Waterevaporates from the wicking surfaces and is carried away by the cool dryair (which has been pre-cooled by the initial passage through the cooldry channel), extracting latent heat, cooling the plates. After passingthrough the wet passage 2, the now warmed moisture-laden return airstream 8 is vented to the atmosphere. In turn, since the pre-cooledreturn air is more effective in lowering the temperature of thehydrophilic surface 7, the incoming air stream 4 in the dry passages 3is more effectively cooled by contact with the dry-side surfaces of theplates 1. The temperature gradient between the air streams on eitherside of the thin thermally conductive plate 1 drives the heat flow fromdry passage side to wet passage side.

The spacing between the plates 1 is approximately the same whether theplates 1 are defining a wet passage 2 or a dry passage 3. In theillustrated example, the wicking material extends about one quarter ofthe way into the wet passages 2, leaving about 50% of the passageunobstructed. Since the volume of return air is only about 50% of theincoming air, it follows that the effective height of the wet airpassage 2 for the return air need only be about 50% of that of theheight of the dry air passage 3 for the incoming air.

The incoming air 4 may be either fresh air from the environment outsidethe living space, or may be recirculation air drawn from inside theliving space.

The present invention is based on the recognition that practicalevaporative cooling can be accomplished in a smaller system when threeconditions are met:

-   -   passage height is reduced,    -   passage length is reduced, and    -   air flow is laminar and at raised shear rates.

While not wanting to be bound by any particular theory of the invention,it is believed that when the plate spacing defining the dry or wetpassages is reduced to approximately 0.6-2.5 mm, preferably 0.6-2.0 mm,more preferably 0.7-1.4 mm, most preferably approximately 0.8-1.2 mm,air can be moved in the passage with laminar flow and high shear rate.With smaller spacing according to the present invention the speedprofile is steep, air flow remains laminar, the boundary layer becomesthinner, and hence heat transfer is dramatically increased. It is knownthat the longer the passage, the greater the pressure drop between inletand outlet. Since a smaller passage length is required in the presentinvention to accomplish the system-effective heat transfer, pressuredrop is kept to a minimum. Compared to the prior art heat exchangersrelying on turbulent air flow to break up the boundary layer, withturbulent air taking energy and increasing back pressure, the inventivedesign actually provides efficiencies in heat transfer as well as airflow in both the wet and the dry passages.

In the case of plate spacings greater than 2.5 mm, preferably 2.0 mm,even if flow is kept laminar, the efficiency of conductive and radiativeheat transfer is lost. With spacings narrower than 0.6 mm air flowbecomes constricted, pressure drop becomes high, and the benefits of theinvention are lost.

By designing a heat exchanger on the basis of this new insight, itbecomes possible to reduce the length of a heat exchanger passages toabout 20% of that of a conventional 400 to 800 mm heat exchanger, i.e.,to about 80 to 200 mm. Thus, in accordance with the invention, it hasbecome possible for the first time to provide a truly compact indirectevaporative cooler.

To graphically illustrate the relationship between various passageheights (H) and passage lengths (L), CFD was carried out using the modelillustrated in FIG. 2. Inlet air was set at 35° C. in all cases. Air wasflowed at 3 m/s between two dry plates gradually cooled to 15° C., thusthe target temperature for exit air is 15° C.

The results are graphically represented in FIG. 3 (showing the resultsfor 0.6 mm to 1.5 mm plate spacing vs. 25 to 100 mm passage length) andFIG. 4 (showing the results for 0.6 mm to 5.0 mm plate spacing vs. 25 to500 mm passage length).

The criteria for selecting feasible combinations of passage height andpassage length are pressure drop and temperature drop, i.e., the airpassage must exhibit (a) “manageable” pressure drop and (b) sufficientcooling. If pressure drop is too large, additional or more powerful fansmay be required to move air, requiring more energy, reducing systemefficiency. Thus, pressure drop should be less than 100 pascal (Pa),preferably less than 60 Pa, more preferably less than 50 Pa, mostpreferably less than 45 Pa. Regarding cooling, air temperature at thepassage outlet should be within 4° C. of target, preferably within 1° C.of target, most preferably within 0.5° C. of target.

In general, it is readily apparent that when passage height is small,pressure drop increases rapidly as passage length is increased. In FIG.3, the pressure drop (Pd) varies from an unacceptable 200 Pa at 0.6 mmpassage height and 100 mm passage length to a negligible 10 Pa at 1.5 mmpassage height and 25 cm passage length.

Considering in greater detail the specific case of the 25 mm passagelength, it can be seen in FIG. 3 that when the passage height is 0.6 mmand source air introduced into the dry passage inlet at 35° C., the airis about 15.3° C. at the passage outlet, thus, heat transfer, measuredas outlet temperature, is quite good. However, as the passage height isincreased, heat transfer efficiency is reduced and the temperature atthe outlet rapidly increases. At a passage height of 0.8 mm the outlettemperature is 16.5° C.; at a passage height of 1.0 mm the outlettemperature is 18.3° C.; at a passage height of 1.2 mm the outlettemperature is 20° C.; at a passage height of 1.5 mm the outlettemperature is 22.5° C.; at a passage height of 2.0 mm (FIG. 4) theoutlet temperature is 25° C.; and at a passage height of 2.5 mm theoutlet temperature is 26.8° C. Evaluating these results for acceptableheat transfer efficiency, according to which air temperature at thepassage outlet should be within 4° C. of target, preferably within 1° C.of target, most preferably within 0.5° C. of target, it follows thatwhen the passage length is 25 mm, the passage height should be no morethan 1.1 mm, preferably no more than 0.7 mm, most preferably not morethan 0.65 mm in height.

Evaluating the same 25 mm passage length at various heights on the basisof pressure drop, it can be seen from FIG. 3 that when the passage is0.6 mm, the pressure drop is an acceptable 50 Pa, and as passage heightincreases, pressure drop decreases further. Thus, for a 25 mm passagelength, pressure drop is acceptable to excellent in all cases.

The remaining passage lengths and heights are evaluated for heattransfer efficiency and pressure drop, and the useful and preferredvalues are collected in Table 1.

TABLE 1 Passage Heat Transfer Pd Length Preferred H Acceptable HPreferred H Acceptable H 500 0.6-4.5 0.6-7.5 ≥2.8 ≥1.9 400 0.6-3.70.6-5.5 ≥2.5 ≥1.7 300 0.6-3.3 0.6-4.5 ≥2.2 ≥1.55 200 0.6-2.5 0.6-3.5≥1.8 ≥1.3 100 0.6-2.2 0.6-2.7 ≥1.5 ≥0.9 75 0.6-1.7 0.6-2.3 ≥1.1 ≥0.8 500.6-1.1 0.6-1.7 ≥0.9 ≥0.7 25 0.6-0.7 0.6-1.1 ≥0.65 ≥0.5

The data in Table 1 is plotted in FIG. 5, wherein dots represent thebounds of heat transfer, Xs represent the bounds of pressure drop, dashlines represent acceptable values, and solid lines represent preferredvalues. The scale of the graph changes above 2.0 mm passage height. Forheat transfer, all passage heights to the left of the lines areincluded. For pressure drop, all passage heights to the right of thelines are included. Since it is necessary for the heat exchanger core toexhibit both good heat transfer and acceptable pressure drop, it followsthat everything between the left and right dash lines is consideredwithin the bounds of the present invention. Everything between the solidlines is particularly preferred core design parameters within the scopeof the present invention.

While the CFD has been carried out only for dry plate spacing andpassage lengths, the addition of wet side to the simulation would nothave been expected to alter the results. While air flowed was a constant3 m/s in the CFD model, and while this flow rate is believed to mostclosely approximate practical conditions, the invention is obviously notlimited to this flow rate.

The heat exchanger of the present invention, with micro-passages asdefined herein, can be adapted to virtually any design, and be used invirtually any system, and offer the benefits of reduced bulk. Generalsystems and materials will be described in the following, without theinvention being limited to the exemplary materials.

The air streams may be in counter-current flowing in oppositedirections, or one stream may be perpendicular to the other stream. Thecompact heat exchanger of the present invention can be used not only forevaporative cooling, but also for heat recovery as discussed in greaterdetail below.

The major difference between the present invention and the prior art isin the reliance on laminar air flow with high shear in narrow airpassages. Thus, prior art structures and methods intended to introduceturbulence into the air flow are to be avoided. For example, in apreferred embodiment of the invention, the heat exchanger plates areformed of a hydrophobic material. One side of the plate remainshydrophobic, while the other side of the plate is rendered hydrophilicby suitable treatment discussed below, and wherein the hydrophilic sideis surface treated, either before or after being rendered hydrophilic,in order to provide sufficient microcapillary or wicking surface forsupply of water for evaporation along the length of the passage. Such a“wet” surface provides very little interference to laminar air flow.

Device for Evaporative Cooling

FIG. 7 shows a sectional view of a practical arrangement for a deviceexploiting the advantages of indirect evaporative cooling. Air entersfrom external ambient through fan 9, which supplies high pressure air tochamber 10. Heat exchanger 11, also referred to as a core, is manifoldedsuch that high pressure air from chamber 10 can only flow through thedry channels of the heat exchanger, and air which flows through the drychannels must flow all the way through the dry channels, emerging intochamber 12. A portion of the air emerging from the dry channels intochamber 12 is required to be turned around to flow back to through thewet channels spaced between the dry channels of the heat exchanger 11.This requires a pressure in chamber 10 to be sufficient to overcome theflow resistance of the wet channels to leave exhaust 13 at the requiredflow rate. This pressure may be achieved by providing a baffle orrestriction in chamber 12 via air duct 14, the pressure differentialacross the baffle at the required flow rate resulting in a staticpressure in chamber 10.

Fan 9 is required to pressurize air to overcome the pressure lossassociated with passing all of the air supplied through the drychannels, plus the static pressure in chamber 12. The static pressure inchamber 12 is sufficient to overcome the flow resistance of theproportion of air flowing through the wet channels to exhaust 13. Thestatic pressure in chamber 12 is regulated by adjusting baffle therebyproducing a static pressure differential across the baffle. The air flowthrough the baffle at such a differential pressure represents a loss ofpower equal to the product of the air flow and pressure differential.This loss is an additional power load on fan 9 which provides noadditional cooling or otherwise useful energy to the air flow. Althoughfan 9 is shown schematically as an axial flow fan, in practice acentrifugal or combined flow fan is generally used due to the highpressures required.

Fan 15 in the exhaust duct 13 of the heat exchanger may optionally beprovided to produce a negative pressure relative to the pressure inchamber 12 sufficient to produce the required air flow through the wetpassages of the heat exchanger. Thus the static pressure immediatelybefore fan 15 will be the sum of the static pressure in chamber 12 andthe pressure differential required for the air flow through the wetpassages of the heat exchanger 11. The operation of the fans 9 and 15can be controlled through electronic speed controllers or other means toproduce a desired ratio of air flow between the supply air and exhaustair. Furthermore, the magnitude and/or ratio of these air flows can bereadily adjusted by varying the speeds of the two fans 9, 15 therebyenabling optimization of the performance of the indirect evaporativecooler. This allows the indirect cooler to operate under a wide range ofconditions through direct control of the fans without the need tointervene and adjust mechanical baffles as in prior art designs, andalso allows for automatic control of the operation of the indirectcooler, for example, under the control of a programmable electroniccontroller.

In a typical indirect evaporative cooler, with a supply to exhaust ratioof 1:1, the fan is required to deliver air at around 600 Pa. If thesupply air required is, say, n units, the power required will be600×2n=1,200n power units. This typically produces a static pressure ofaround 150 Pa in chamber 48 and thus a pressure differential of 150 Paacross the wet passages to exhaust. The pressure differential across thedry passages of the heat exchanger is 600 Pa-150 Pa=450 Pa.

The Plates

The plates 16 separate dry product air passages and wet working airpassages. Each plate is made of a thin material to allow easy heattransfer across this plate and thus to readily allow heat to transferfrom the dry product passage to the wet working passage. The plate ispreferably formed of a plastic film such as a thin-wall dense film orsheet of polyethylene, polypropylene, polystyrene, polyvinyl chloride,polyethylene terephthalate, or the like material having good vaporbarrier properties. Although a low permeability material such as plasticdoes not readily transfer heat, heat transfer perpendicularly throughthe plastic layer will be good since the plates according to the presentinvention are very thin.

Air streams generally flow between two plates rather than across oneplate. If two such plates are aligned with dry sides facing, then theair streams flow between the two plates on the dry sides; if the wetsides are facing, the air streams flow between the plates on the wetsides. In embodiments having more than two plates, air streams may firstflow between the dry sides of two plates, then flow through one or bothplates and enter wet passages, in which they will flow across one of thetwo previous plates (on the reverse side) and the wet side of a thirdplate.

Plates of various materials and configurations are described in U.S.Pat. Nos. 6,581,402; 6,705,096; 7,228,699; and 8,468,846, thedisclosures of which being incorporated herein by reference.

The Wicking Material/Surface Treatment

While the plate 16 may be formed of a hydrophobic polymer such asextruded polypropylene, one surface of the polymer may be renderedsubstantially hydrophilic by subjecting it to Corona treatment, plasmadischarge, plasma jet, flame treatment, acid etching and nano-surfacingor nano-coating.

The hydrophilic side of the plate may be provided with an additionalliquid retaining layer formed from a fibrous non-woven material 17.Although reference may be made to a liquid retaining surface, it isclearly understood that the surface is in fact a liquid retaining andreleasing surface. The evaporation rate off of a hydrophobic woven orspun bond material where water has been impregnated in between thefibers is higher than from a hydrophilic material where water has beenabsorbed into the material and between the fibers. This means that amuch smaller temperature difference across the plate is required toachieve the same evaporation rate, which therefore increases the heattransfer rate. See for example WO 2010/011687 (Gillan) teaching ahydrophobic fiber sheet formed to wick evaporative fluid.

In the present invention the wicking material or wicking layer/treatmentis relatively thin, and thus contains only a small amount of water.Accordingly, the system is efficient in that air rather than water iscooled.

The wicking material may be hydrophilic or hydrophobic, and suitablematerials include cellulose, fiberglass, organic fibers, organic-basedfibers, porous plastics, carbon-based fibers, polyesters, polypropylene,silicon-based fibers and combinations of these substances. The wicklayer material may be in a number of forms: films, weaves, braids,fibers, beds of particles such as beads and combinations thereof.

A substantially compliant nonwoven wicking material is disposed on andfixedly attached at a number of locations to the hydrophilic surface ofthe first polymer substrate. Similarly, substantially compliant nonwovenmaterial is also disposed on and fixedly attached at a number oflocations to the hydrophilic surface of the second polymer substrate.The substantially compliant nonwoven material can be a spunbondedmaterial, a melt blown material, hydroentangled (spunlaced) material ormade through any other processes such as co-forming, airlaying,wetlaying, carding webs, thermal bonding, needle punching, chemicallybonding or combinations thereof. Embodiments of spunbonded materialinclude polyolefin, Polyethylene terephthalate (PET) and nylon.Embodiments of melt blown material include polyolefin, Polyethyleneterephthalate (PET) and nylon. Embodiments of hydroentangled materialinclude cotton, rayon or viscose staple fiber, lyocell staple fiber,polyolefin staple fiber, polyester staple fiber and nylon staple fiber.

Nonwoven webs can be formed from fibers and filaments based onhydrophobic or hydrophilic polymers. Representative, but not complete,examples of polymers that are hydrophobic for making nonwoven webs arepolyolefins and polyethylene terephthalate. Representative, but notcomplete, examples of hydrophilic polymers for making nonwoven websinclude cellulosic materials like cotton, rayon or viscose etc. Theapplication of the fact that under suitable conditions of porosity,fiber/filament diameter, density (GSM) etc, significant capillary actionand wicking of water can occur in a web has been innovatively applied inthe invention. The invention innovatively utilizes the porosity ofcertain porous nonwoven webs that can often be sufficient to enable theeasy transport of water and other fluids because of wicking caused bycapillary action.

The prior art teaches that hydrophilic materials can better hold water.However, in relation to the cooling apparatus, the application of thisquality has a disadvantage that in the case of a nonwoven web made fromhydrophilic polymers, some of the water will swell the fibers and therest will go around and over the fibers. This would lose the rigidity inthe heat exchanger pads. Further, in relation to the cooling apparatus,hydrophilic non-woven would swell, while one of our objectives is toretain the thinnest film of the water to facilitate better heat transferand evaporation. Porous low density nonwoven webs made from hydrophobicfibers or filaments can transfer water through wicking action. Water canflow along, around and over but not through the hydrophobic polymerfibers. The porosity and associated wicking action by a porous nonwovenweb can render the nonwoven web effectively hydrophilic in terms of itscapability to be wet and easily spread water even if the fibers orfilaments constituting the nonwoven web are made from hydrophobicpolymers. The invention thus innovatively employs the materials known tobe hydrophobic for the retention of water as required. The inventionovercomes the problem in maintaining rigidity of heat exchanger pads dueto the use of hydrophilic material, as evidenced by relevant prior art,by employing hydrophobic material.

Examples of fibers that are hydrophobic are polyolefins and polyethyleneterephthalate. Porous low density nonwoven webs made from thesehydrophobic fibers or filaments can be hydrophilic through wickingaction.

Spacers

The plates may be separated by any conceivable spacing means 18. Theplates can be deformed with a punch or a roller to introduce raisedpoints or walls. Adhesive or plastic can be printed to provide thedesired structures at the desired height.

The spacers may be spaced apart a distance of 30-50 mm. Keeping in mindthe thinness and minimal structural integrity of the plates, the spacersshould be spaced as far apart as possible while still ensuring properspacing of the sheets.

Heat Scavenging Heat Exchanger

The device may be operated in winter months to scavenge heat fromexhaust gases of a space and thus pre-heat fresh air, whilesimultaneously humidifying the fresh air.

Such an indirect evaporative cooler would have cycle selection means, sothat during summer months, it may be used to provide cooled,non-humidified air, and during winter months, it may be used to scavengeheat from gases exiting a space while simultaneously humidifying thespace.

In winter months, it is advantageous to exchange heat between exhaustair leaving a warmed space and cold fresh air being brought in from theatmosphere, i.e. the outdoor air or other source of environmental air.This reduces the heat required to warm the fresh air. The presentinvention also allows the addition of humidity to the fresh air, thusaddressing another winter problem: cold outside air that has condensedmoisture out and therefore has a low absolute humidity or extremely dryair that results in dry inside air as the moisture on the inside reduceswith fresh air changes with the outside. The “cycle selection” as towhich stream of air is exhausted to the atmosphere, and which goes tothe space to be conditioned, is a feature of embodiments having thisarrangement.

Modular Design/Scaling Up

Heat exchangers can be manufactured in standard sizes as modules, and anumber of these modules can be assembled to provide an air handler witha capacity required for a given living or working space.

System Operation

During operation, an evaporative liquid is distributed to thesubstantially compliant nonwoven material of each of the hydrophilicsurfaces of the unit and a fluid flows within a space separating each ofthe units with heat being exchanged between the evaporative liquid andthe fluid, and another fluid flowing through at least some passages fromthe number of passages. The other fluid exchanges heat with thehydrophilic surface. (Heat is transferred from the nonwoven material tothe hydrophilic surface/wet side and through the plate to thehydrophobic surface/dry side). In one instance, the fluids are air, andthe heat exchanger has moist air from the evaporated cooling of thesubstantially compliant nonwoven material having the evaporative liquiddistributed over and dry air flowing through the passages and beingcooled. In that instance to a heat exchanger is referred to as a DryAir, Moist Air (DAMA) heat exchanger.

As disclosed hereinabove, in the indirect evaporative cooling heatexchanger of this invention described herein above, during operation, anevaporative liquid is distributed to the substantially compliantnonwoven material of each of the hydrophilic surfaces of a unit and asecondary fluid flows within a space separating each of the units withheat being exchanged between the evaporative liquid and the fluid. Aprimary fluid flows through at least some passages from a number ofpassages. The primary fluid exchanges heat with the hydrophilic surface.(Heat is transferred from the nonwoven material to the hydrophilicsurface and through the substrate to the hydrophobic surface). Anotherevaporative liquid is distributed to the direct evaporative component.The direct evaporative component is positioned downstream from theindirect evaporative cooling component and receives at least a portionof the primary fluid from the indirect evaporative cooling component. Inthe direct adiabatic evaporative component, the primary fluid is cooledby addition of the other evaporative liquid. During the cooling processin the adiabatic direct evaporative component, there is no change in thetotal energy or enthalpy, but, a portion of sensible heat of the primaryfluid is converted into latent heat.

In one instance, the two-stage evaporative cooling apparatus of thisinvention includes a primary fluid supply component, located upstreamfrom the indirect evaporative cooling component of this invention andsupplying the primary fluid to the indirect evaporative coolingcomponent of this invention. During operation the primary fluid supplycomponent draws ambient fluid through a filtering component and suppliesfiltered ambient fluid as the primary fluid. The filtering component caninclude one of a variety of filters (including but not limited toconventional filters, carbon filters, electrostatic filters, etc.). Afirst evaporative liquid supply system supplies the evaporative liquidto the indirect evaporative cooling component of this invention. Asecond evaporative liquid supply system supplies the other evaporativeliquid to the direct adiabatic evaporative cooling component. In oneinstance, a liquid holding component (such as a tank) provides a supplyof the evaporative liquid and the other evaporative liquid. The firstevaporative liquid supply system (a pump in one instance) and the secondevaporative liquid supply system (another pump in one instance) aredisposed inside the liquid holding component. The first evaporativeliquid supply system, the second evaporative liquid supply system and aliquid holding component are comprised of aseptic material. A liquiddisinfection system can be disposed to receive the evaporative liquidand the other evaporative liquid and render both of them disinfected. Inone embodiment the liquid disinfection system includes a systemutilizing ultraviolet (UV) radiation in order to disinfect theevaporative liquid on the other evaporative liquid. It should be notedthat other liquid disinfecting systems, such as, but not limited to,system utilizing ozone and other liquid disinfecting systems are withinthe scope of this invention.

Note that the indirect evaporative cooler of the invention may alsoaccomplish both direct and indirect evaporative cooling of the productair stream. A portion of the dry sides may be wetted, in a mannersimilar to the wick materials used on wet sides or in a differentmanner, so as to cause further cooling of the product air stream. Thiswet portion of the dry sides may advantageously be placed downstream ofthe dry portion of the dry sides, so that prior to being humidified inthe direct evaporation cooling process, the sensible temperature of theproduct air stream is reduced as much as possible. One particularadvantage of this ordering is that below approximately 65 degrees F.,modest increases in humidity cause disproportionate reduction in airtemperatures, in accordance with standard psychometric charts. Inanother embodiment of the invention, this wetted portion of the drysides constitutes the final 1 to 25 percent of the surface area of thedry passages.

Some embodiments of the invention may utilise 3D printing forconstruction of the device. Accordingly, in some embodiments the presentinvention may reside in a digital blueprint comprising a digital file ina format configured for use with rapid prototyping and computer aideddesign (CAD) and/or manufacturing, such as being in the STL(stereolithography) file format. Such digital blueprint files, whetherproduced by performing a three dimensional scan of an embodiment of theinvention, or produced by a CAD development software tool, or the like,are within the scope of the present invention.

It will be appreciated by persons skilled in the art that numerousvariations and/or modifications may be made to the invention as shown inthe specific embodiments without departing from the spirit or scope ofthe invention as broadly described. The present embodiments are,therefore, to be considered in all respects as illustrative and notrestrictive.

1. An indirect evaporative cooler comprising: a heat exchanger corehaving heat exchange plates defining alternating wet and dry air flowpassages, wherein the relationship between the passage height, measuredas space between plates defining at least the dry air flow passages, andthe length of the air flow passages falls within an area on a graphdefined by the following points, wherein passage length is plotted onone axis and passage height, measured as distance between plates, isplotted on another axis: Passage Passage Length (mm) Height (mm) 5001.9-7.5 400 1.7-5.5 300 1.55-4.5  200 1.3-3.5 100 0.9-2.7 75 0.8-2.3 500.7-1.7 25  0.6-1.1.


2. The indirect evaporative cooler as claimed in claim 1, wherein therelationship between the passage height of at least the dry air flowpassages and the length of the passages falls within an area on a graphdefined by the following points, wherein passage length is plotted onone axis and passage height is plotted on another axis: Passage PassageLength (mm) Height (mm) 500 2.8-4.5 400 2.5-3.7 300 2.2-3.3 200 1.8-2.5100 1.5-2.2 75 1.1-1.7 50 0.9-1.1 25 0.65-0.7. 


3. The indirect evaporative cooler as claimed in claim 1, whereinpassage length is from 80 to 200 mm.
 4. The indirect evaporative cooleras claimed in claim 1, wherein the space between plates defining drypassage height is from 0.6 mm to 2.5 mm.
 5. The indirect evaporativecooler as claimed in claim 4, wherein the space between plates definingdry passage height is from 0.7 mm to 1.4 mm.
 6. The indirect evaporativecooler as claimed in claim 5, wherein the space between plates definingdry passage height is from 0.8 mm to 1.2 mm.
 7. The indirect evaporativecooler as claimed in claim 1, wherein the space between plates definingwet passage height is from 0.6 mm to 2.5 mm.
 8. The indirect evaporativecooler as claimed in claim 7, wherein the space between plates definingwet passage height is from 0.7 mm to 1.4 mm.
 9. The indirect evaporativecooler as claimed in claim 8, wherein the space between plates definingwet passage height is from 0.8 mm to 1.2 mm.
 10. A method of indirectevaporative cooling, the method comprising directing a laminar flow ofair at a flow rate of 2.5-7.0 m/s through at least dry air passages of aheat exchanger having heat exchange plates defining alternating wet anddry air flow passages, the plates having a separation of from 0.6 mm to2.0 mm and defining air passages having a length of from 25 to 300 mm,dividing the air after passing through the dry passages into first andsecond air streams, directing the first air stream into the wet air flowpassages in counter-current flow to the airflow in the dry passages, anddirecting the second air stream to a space to be cooled.
 11. The methodof indirect evaporative cooling as claimed in claim 10, wherein air isflowed through the dry air flow passages at 3.0-7.0 m/s.
 12. The methodof indirect evaporative cooling as claimed in claim 10, wherein air isflowed through the wet air flow passages at 2.5-4.0 m/s.
 13. An indirectevaporative cooler comprising: a heat exchanger core having heatexchange plates defining a plurality of wet air flow passages and aplurality of dry air flow passages; and at least one fan configured todrive air through the dry air passages, the dry air passages beingconfigured so that a substantially laminar airflow having a raised shearrate arises in the dry air passages, and so that a back pressure acrossa length of the dry air passages remains acceptably low.
 14. Theindirect evaporative cooler as claimed in claim 2, wherein passagelength is from 80 to 200 mm.
 15. The indirect evaporative cooler asclaimed in claim 14, wherein the space between plates defining drypassage height is from 0.6 mm to 2.5 mm.
 16. The indirect evaporativecooler as claimed in claim 5, wherein the space between plates definingwet passage height is from 0.6 mm to 2.5 mm.
 17. The indirectevaporative cooler as claimed in claim 15, wherein the space betweenplates defining wet passage height is from 0.6 mm to 2.5 mm.
 18. Theindirect evaporative cooler as claimed in claim 17, wherein the spacebetween plates defining wet passage height is from 0.7 mm to 1.4 mm. 19.The indirect evaporative cooler as claimed in claim 18, wherein thespace between plates defining wet passage height is from 0.8 mm to 1.2mm.
 20. The method of indirect evaporative cooling as claimed in claim11, wherein air is flowed through the wet air flow passages at 2.5-4.0m/s.